1. Field of the Invention
The present invention relates to turbomachineries such as pumps for transporting liquids or compressors for compressing gases, and relates in particular to turbomachineries comprising an impeller having short splitter blades between full blades for improving performance.
2. Description of the Related Art
FIGS. 1(a)-1(c) show a normal impeller comprised only by full blades. This type of impeller has a plurality of blades 3 on a curved outer surface of a truncated cone shaped hub 2 disposed equidistantly along a circumferential direction around a shaft 1. Flow passages are formed by a space formed by a shroud (not shown), two adjacent blades and the curved hub surface. The fluid enters the impeller space through an inlet opening near the shaft and flows out through the exit opening at the outer periphery of the impeller. The fluid is compressed and given a kinetic energy by the rotational motion of the impeller about the shaft so as to enable pressurized transport of the fluid by the turbomachinery.
Although some impellers are unshrouded, the clearance between the casing and the blade tip is set minimal so as to prevent a leakage flow therefrom. Therefore, the flow within the unshrouded impeller is substantially the same as that of an impeller having a shroud. Thus, in the explanations given for impellers having a shroud in this specification hereinafter, a term xe2x80x9cshroud-sidexe2x80x9d should be construed as xe2x80x9ccasing sidexe2x80x9d or xe2x80x9cblade tip sidexe2x80x9d for the unshrouded impellers.
One of the significant problems to be solved for such conventional turbomachineries is not only to improve their performance at a design flow rate, but to realize a wide operating range. For example, when pumps are operated at a flow rate beyond the design flow rate, local increase in the fluid velocity induces a local pressure drop at an inlet region of the impeller. And when the suction pressure is low, in particular, the fluid pressure will become less than the vapor pressure of the fluid in some regions. The result is a generation of so-called xe2x80x9ccavitationxe2x80x9d in which the fluid is vaporized, and it is well known that a pressurization effect of the pump is deteriorated due to blockage effect of bubbles.
On the other hand, if a compressor for compressing gas is operated at a flow rate beyond the design flow rate, the velocity becomes higher than the acoustic velocity in a region of the minimum cross section of the flow passage to cause a phenomenon of so-called xe2x80x9cchokingxe2x80x9d, and it is well known that, due to blocking of the gas passage, a compressing effect of the compressor is rapidly lost.
Such problems of degradation in the device performance, due to cavitation and choking phenomena, are caused by the fact that the pressurizing action of the impeller is interrupted due to reduction of the effective flow passage area, which is brought about by the enlargement of the vaporization regions for liquids or supersonic velocity regions for gases. An effective solution for improving suction capability of the turbomachinery is, therefore, to enlarge the flow passage area at an inlet region of the impeller. One approach is to remove a fore part of every other blade. In this case, those blades having the original blade length are called xe2x80x9cfull bladesxe2x80x9d and those with shorter blade length are called xe2x80x9csplitter bladesxe2x80x9d. Such impellers having splitter blades aim to increase the suction capability by increasing the flow passage area at an inlet region of the impeller by reducing the effective number of blades, and at the same time, the pressurizing effect of the blades is maintained in the latter part of the flow passage by splitter blades placed between the full blades.
FIGS. 2A-2C illustrate a conventional impeller with splitter blades. The impeller comprises full blades 4 and splitter blades 5 alternatingly on the hub 2 so that it can secure a wide flow passage at the inlet, and in the latter half, sufficient number of blades are provided to secure adequate pressurization effects. As described above, in view of convenience for manufacturing, such splitter-bladed impellers are made by machining off the fore part of every other full blade disposed equidistantly around the hub. The shape of the splitter blade is identical to that of the full blade except for the removed region, and the splitter blades are placed at the mid-pitch locations between the full blades.
However, in such an impeller having splitter blades made by removing a fore part of every other evenly spaced full blade, the fluid velocity at the suction surface 4s of a full blade 4 facing the inlet opening is increased while the fluid velocity at the pressure surface 4p of the opposite full blade 4 is decreased. Under these conditions, in the fore part of the flow passage where the leading half of the full blade is removed, the fluid cannot flow right in the direction along the blade surfaces. The result is a generation of flow fields mismatch due to the difference in the fluid flow angles and the blade angles at the inlet of the splitter blade, which induces a problem of flow separation at the splitter blade.
FIG. 3A shows a meridional geometry of the impeller with splitter blades shown in FIGS. 2A-2C having a specific speed of 400 (m3/min,m,rpm), and FIG. 3B is a contour diagram of meridional velocities of the flow on a ring-shaped flow passage formed at a section Axe2x80x94A in FIG. 3A, computed by a three-dimensional viscous flow calculation. FIGS. 4A-4B show similar diagrams for the impeller having a specific speed of 800 (m3/min,m,rpm). As can be understood from these drawings, the fluid velocities on the suction-side of the full blade are significantly higher over the area from the hub to the shroud than those on the pressure side, so that the mass of fluid passing through the impeller becomes more concentrated on the suction-side of the full blade.
When the splitter blade is positioned at a mid-pitch location between the full blades under such flow conditions, a phenomenon of flow imbalance is generated such that the mass of fluid flowing in the flow passage formed between the suction surface 4s and the pressure surface 5p is different from that between the pressure surface 4p and the suction surface 5s. This produces a disparity in such fluid dynamic parameters as outflow velocity and outflow angle at both sides of every splitter blade. It is known that such disparities cause a number of undesirable effects such as an increased loss due to flow mixing downstream of the impeller, and lowering of performance in the downstream diffuser section due to increased unsteadiness of the outflow from the impeller.
To relieve such mismatching in flow fields and non-uniformity in the flow passage for improving the performance of the impeller, it is generally considered that the splitter blade leading edge should be moved from the mid-pitch location towards the suction-side of the adjacent full blade. FR-A-2550585 is an example of teaching in this regard. For example, some of the remedial approaches to flow rate mismatching include: to reduce mismatching at the fluid inlet by making the flow passage width sizes the same on both sides at the splitter blade leading edge; to reduce the detrimental effect of flow rate non-uniformity by making the splitter blade trailing edge to be located at the same distance ratio between the full blades as its leading edge; and to displace the circumferential location of the splitter blades for optimizing the flow rate.
However, such known remedial techniques are not satisfactory enough to adequately optimize the position of the splitter blades. Specifically, as seen in FIGS. 3A, 3B, 4A and 4B, pitchwise or circumferential expansion of the high velocity region varies non-uniformity of the flow rate changes radically between the hub-side and shroud-side of the flow passage. Also, the fluid velocity is especially high on the shroud-side of the suction surface of the full blade, where flow rate inhomogeneity in the spanwise direction is also generated. Therefore, because the conventional techniques do not consider the effects of the three-dimensional nature of the fluid velocity distribution, adverse effects of the flow rate inhomogeneity on device performance have not been fully eliminated.
It is an object of the present invention to solve the problems of depressed performance caused by improper shape of the splitter blade and provide a clear design of proper splitter blades so as to provide an impeller with splitter blades having a wide operating range without affecting the performance of the turbomachinery.
The object has been achieved in an impeller for a turbomachinery comprising: a hub; a plurality of full blades equidistantly disposed on the hub in a circumferential direction; and a plurality of splitter blades disposed between each adjacent two of the full blades, wherein each of the splitter blades is shaped in such a way that a spanwise distribution of a pitchwise position of a leading edge of the splitter blade is determined according to a spanwise and pitchwise non-uniformity distribution of fluid velocity of a fluid flowing into the splitter blade, as illustrate by a schematic drawing shown in FIG. 5. Here, the term xe2x80x9cspanwisexe2x80x9d is used for a xe2x80x9cthicknessxe2x80x9d direction of the impeller, that is, a direction along a straight line tying two corresponding points on the hub and the shroud (blade tip) in a meridional cross section as shown in FIGS. 3A or 4A. Also, the term xe2x80x9cpitchwisexe2x80x9d is used for a circumferential direction within a pitch between two adjacent full blades as shown in FIGS. 5A and 5B.
By adjusting the position of the splitter blade leading edge in the hub-to-shroud space, the impeller of the present invention with splitter blades enables mismatching of flow fields or non-uniform flow rates in the flow passages to be prevented, as well as the onset of impeller stall in partial flow regions to be prevented or destroyed. Therefore, it is possible to moderate the adverse effects of three-dimensional non-uniformity in the flowfields in the hub-to-shroud space in the impeller, so as to provide a high efficiency operation of the turbomachinery.
Each of a flow passage formed between the full blade and the splitter blade may be shaped in such a way that a flow separation on the aft part of the suction surfaces of the full blade and the splitter blade is avoided.
Also, each of the splitter blades may be shaped in such a way that a position of a leading edge of the splitter blade at a blade tip is displaced away from a mid-pitch position of adjacent full blades, and the leading edge of each of the splitter blades has a predetermined distribution of pitchwise position varying along a spanwise direction.
The distribution of the circumferential position may be determined according to a non-uniformity distribution of fluid flowing into the splitter blade.
It is desirable to locate any position of the leading edge within a range of non-dimensional parameter P as expressed in an inequality relation: 0.42 less than P less than 0.77, where P is a pitchwise distance between the position and a circumferentially corresponding position on a blade camber line of a full blade adjacent to a suction side of the splitter blade which is normalized by a pitch distance between adjacent full blades (refer to FIG. 6).